Engine control method and apparatus

ABSTRACT

An engine is designed to allow a compression self-ignition combustion under an air-fuel ratio leaner than a stoichiometric air-fuel ratio to be performed at least in a partial-load range of the engine. Under a condition that an engine speed varies at a same load in an engine operating region of the compression self-ignition combustion, a compression end temperature Tx, which is an in-cylinder temperature just before an air-fuel mixture self-ignites, is controlled to be raised higher in a higher engine speed side than in a lower engine speed side. As one example of control for the compression end temperature Tx, an internal EGR amount is controlled to be increased larger in the higher engine speed side than in the lower engine speed side, to raise a compression initial temperature T 0  which is an in-cylinder temperature at a start timing of a compression stroke. This makes it possible to perform the compression self-ignition combustion under a lean air-fuel ratio in a wider engine speed range to effectively enhance engine thermal efficiency.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an engine control method and apparatusdesigned to allow a compression self-ignition combustion to be performedin at least a part of an engine operating region.

2. Description of the Background Art

Heretofore, in a spark-ignition gasoline engine having a spark plug, amethod has been implemented which is designed to perform a compressionself-ignition combustion based on self-ignition of an air-fuel mixture,in a partial engine load operating region, and to perform aspark-ignition (SI) combustion based on forced ignition of an air-fuelmixture by a spark plug, in another operating region on a higher engineload side than the partial engine load operating region, as disclosed inJP 2007-292060A.

As for the method designed to selectively use the compressionself-ignition combustion and the spark-ignition (SI) combustiondepending on engine load in the above manner, various techniques haveheretofore been proposed. Specifically, the compression self-ignitioncombustion is a combustion mode where an air-fuel mixture self-ignitessimultaneously at many positions in a combustion chamber, and consideredto have a potential to provide higher efficiency than the commonly-usedspark-ignition combustion. However, the compression self-ignitioncombustion has a problem with combustion controllability (i.e.,preignition or knocking is more likely to occur) when the engine load ishigh, so that it is necessary to perform the spark-ignition combustionwith high controllability, in a high engine load range. For thispurpose, the two combustion modes, i.e., the compression self-ignitioncombustion and the spark-ignition combustion, are selectively useddepending on engine load. This allows the compression self-ignitioncombustion to be adequately performed, but partially, so that enginethermal efficiency is enhanced in an engine operating region subject tothe compression self-ignition combustion, which provides an advantage ofbeing able to improve fuel economy performance.

In the engine disclosed in the JP 2007-292060A, the compressionself-ignition combustion and the spark-ignition combustion areselectively used depending on engine speed as well as engine load.Specifically, the engine is designed to perform the compressionself-ignition combustion primarily in a low engine speed range, andswitch the compression self-ignition combustion to the spark-ignitioncombustion, in the remaining engine speed range. The reason forswitching to the spark-ignition combustion in a high engine speed rangeis that, when the engine speed is high, a period where an air-fuelmixture is exposed to high temperatures and high pressures(high-temperature/high-pressure period) becomes relatively short, sothat the air-fuel mixture is less likely to self-ignite, and misfire ismore likely to occurs.

Meanwhile, it is known that, as means for enhancing the engine thermalefficiency, a technique may be employed which is designed to burn anair-fuel mixture under an air-fuel ratio leaner than a stoichiometricair-fuel ratio. Specifically, when air is introduced into the cylinderin an excess amount with respect to a fuel (gasoline) supply amount toperform combustion under a resulting lean air-fuel ratio, a combustiontemperature can be lowered as compared with combustion under thestoichiometric air-fuel ratio, so that an exhaust loss and a coolingloss of the engine are reduced, which makes it possible to furtherenhance the engine thermal efficiency.

Therefore, if the compression self-ignition combustion can be performedunder the condition that the air-fuel ratio is set to a lean value, itis expected to make it possible to more effectively enhance the enginethermal efficiency so as to improve the fuel economy performance.

However, it is assumed that, if the air-fuel ratio is simply enleaned,self-ignitability of an air-fuel mixture will deteriorate, and therebymisfire is more likely to occur when switched to the compressionself-ignition combustion. Consequently, an engine speed range whichallows for the compression self-ignition combustion is narrowed, andthereby the effect of enhancing the engine thermal efficiencydeteriorates. Thus, there is a need for a technique capable of allowingthe compression self-ignition combustion to be performed in a widerengine speed range, even under an enleaned air-fuel ratio.

SUMMARY OF THE INVENTION

In view of the above circumstances, it is an object of the presentinvention to allow a compression self-ignition combustion under a leanair-fuel ratio to be performed in a wider engine speed range so as tofurther effectively enhance engine thermal efficiency.

In order to achieve the above object, the present invention provides amethod for controlling an engine. The method comprises a step ofallowing a compression self-ignition combustion under an air-fuel ratioleaner than a stoichiometric air-fuel ratio to be performed at least ina partial-load range of the engine, wherein, under a condition that anengine speed varies at a same load in an engine operating region of thecompression self-ignition combustion, a compression end temperature,which is an in-cylinder temperature just before an air-fuel mixtureself-ignites, is controlled to be raised higher in a higher engine speedside than in a lower engine speed side.

The present invention also provides an apparatus for controlling anengine. The apparatus comprises a controller adapted to controlrespective sections of the engine to allow a compression self-ignitioncombustion under an air-fuel ratio leaner than a stoichiometric air-fuelratio to be performed at least in a partial-load range of the engine,wherein the controller is operable, under a condition that an enginespeed varies at a same load in an engine operating region of thecompression self-ignition combustion, to control a compression endtemperature, which is an in-cylinder temperature just before an air-fuelmixture self-ignites, in such a manner that it is raised higher in ahigher engine speed side than in a lower engine speed side.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a graph showing a relationship between an excess air ratio λand an NOx generation amount.

FIG. 2 is an explanatory diagram showing a state in which a compressionself-ignition combustion occurs through a chemical reaction between fueland oxygen.

FIG. 3 is a schematic diagram showing a typical reciprocating gasolineengine.

FIG. 4 is a graph prepared by calculating a condition of a compressionend temperature and a compression end pressure for causing an air-fuelmixture to self-ignite at an MBT ignition timing when an engine speed is1000 rpm.

FIG. 5 is a graph prepared by calculating conditions of the compressionend temperature and the compression end pressure for causing theair-fuel mixture to self-ignite at the MBT ignition timing when theengine speed is 1000, 2000, 3000, 4000, 5000 or 6000 rpm.

FIG. 6 is a schematic diagram showing an overall configuration of anengine according to one embodiment of the present invention.

FIG. 7 is a chart showing one example of an intake/exhaust valve liftcharacteristic which is set to control an internal EGR amount.

FIG. 8 is a chart showing one example of an intake valve liftcharacteristic which is set to control an effective compression ratio.

FIG. 9 is an explanatory diagram showing a change in engine operatingstate.

FIG. 10 is an explanatory diagram showing a control scheme to beexecuted when the engine operating state is changed along a engineload-axis direction.

FIG. 11 is an explanatory diagram showing a control scheme to beexecuted when the engine operating state is changed along an enginespeed-axis direction.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

<Researches Before Reaching the Present Invention>

(1-1) Discussion about Air-Fuel Ratio

As means for enhancing engine thermal efficiency, a technique may beemployed which is designed to burn an air-fuel mixture under an air-fuelratio leaner than a stoichiometric air-fuel ratio, as described in the“BACKGROUND OF THE INVENTION”. Based on making the air-fuel ratio lean,a combustion temperature of the air-fuel mixture is lowered, so that anexhaust loss and a cooling loss are reduced to provide enhanced enginethermal efficiency.

A gasoline engine designed to perform such a lean combustion (lean-burngasoline engine) has been developed for years. For example, a gasolineengine designed to be operated using an air-fuel mixture having anair-fuel ratio set to a large value of up to about 20 (by comparison,the stoichiometric air-fuel ratio is generally about 14.7) has been putto practical use in the past. However, even if the air-fuel ratio is setto about 20, it is difficult to expect significant improvement in theengine thermal efficiency, and there is a problem of deterioration inemission performance. Specifically, a three-way catalyst capable ofsimultaneously purifying HC, CO and NOx contained in exhaust gas isgenerally provided in an exhaust passage of an engine. The three-waycatalyst can bring out maximum performance when the air-fuel ratio isequal to the stoichiometric air-fuel ratio. Thus, for example, when theair-fuel ratio is enleaned up to about 20, NOx purifying performancewill significantly deteriorate.

In cases where the NOx purifying performance cannot be expected from thethree-way catalyst, if a NOx catalyst, such as NOx storage/reductioncatalyst, is provided in addition to the three-way catalyst, NOx can beobviously purified to some extent to suppress a NOx emission level. Inthe previously developed lean-burn engine, the NOx problem has beencleared by providing the NOx catalyst in the exhaust passage. However,the performance of the NOx catalyst has limitations. For example, underan air-fuel ratio of about 20, even if the NOx catalyst is additionallyprovided, it becomes difficult to meet emission regulations which arebecoming more stringent year after year.

Under the above circumstances, the inventors of this application came upwith a concept of reducing an amount of NOx itself to be generated bycombustion (raw NOx amount), by drastically enleaning the air-fuel ratiowith respect to the stoichiometric air-fuel ratio. Specifically, it isintended to drastically enlean the air-fuel ratio so as to lower acombustion temperature of an air-fuel mixture to a value less than a NOxgeneration temperature (a temperature at which NOx is activelygenerated) to significantly reduce the NOx generation amount.

FIG. 1 is a graph showing a relationship between an excess air ratio λand the NOx generation amount (combustion-generated raw NOx amount). Inthis graph, the value Y on the vertical axis indicates a NOx amountwhich can sufficiently meet the emission regulations if a NOx catalystis provided, and the value X on the vertical axis indicates a NOx amountwhich can sufficiently meet the emission regulations without providing aNOx catalyst.

As can be seen from FIG. 1, when the excess air ratio λ which is a ratioof an actual air-fuel ratio to the stoichiometric air-fuel ratio (avalue obtained by dividing an actual air-fuel ratio by thestoichiometric air-fuel ratio) is increased to 2.4 (equal to an actualair-fuel ratio of about 35) or more, the combustion-generated raw NOxamount becomes equal to or less than the reference value X, so that itbecomes possible to sufficiently meet the emission regulations withoutproviding a NOx catalyst. When the excess air ratio λ is in the range of2 to less than 2.4, the NOx amount is less than the reference value Yalthough it becomes greater than the reference value X, so that it ispossible to sufficiently meet the emission regulations as long as a NOxcatalyst is provided.

For the above reasons, it is proven that the problem concerning the NOxemission level can be cleared by setting the excess air ratio λ, i.e., aratio of an actual air-fuel ratio to the stoichiometric air-fuel ratio,to 2 or more, more preferably 2.4 or more. When the excess air ratio λis set to 2.4 or more (λ≧2.4), a NOx catalyst can be omitted, which ismore advantageous in terms of cost, than when the excess air ratio λ isset in the range of 2 to less than 2.4.

However, when it is attempted to burn an air-fuel mixture under theultra-lean air-fuel ratio (λ≧2 (or 2.4)), a flame propagation velocityafter ignition of the air-fuel mixture is significantly lowered, ascompared with combustion under the stoichiometric air-fuel ratio. Thus,if the same spark-ignition combustion as that in a conventional gasolineengine is used for this concept, misfire is more likely to occur, andthereby it is difficult to practically realize the concept. In contrast,when a compression self-ignition combustion where an air-fuel mixtureself-ignites simultaneously at many positions, is used for the concept,there is a possibility that an adequate combustion can be performed evenunder the ultra-lean air-fuel ratio (λ≧2), irrespective of the lowing ofthe flame propagation velocity, if only an environment allowingself-ignition can be created. With a focus on this point, the inventorsfurther studied the following factors.

(1-2) Discussion about Compression Self-Ignition Combustion

The compression self-ignition combustion in a gasoline engine is aphenomenon that fuel (gasoline) and oxygen (O₂) chemically react witheach other on an autonomous basis, as shown in FIG. 2. When a chemicalreaction between fuel and oxygen, water and carbon dioxide are producedwhile generating heat due to an internal energy difference. Whether sucha chemical reaction occurs is determined by a temperature and a pressureof an air-fuel mixture, and a high-temperature/high-pressure period(period where the air-fuel mixture is exposed to high temperatures andhigh pressures). More specifically, a molecular velocity becomes higheras the temperature becomes higher, and a molecular collision frequencybecomes higher as the pressure (i.e., molecular density) becomes higher.Thus, as the temperature and the pressure become higher, energyresulting from collision between fuel and oxygen is more increased, andthe chemical reaction is more likely to occur. Then, after the periodwhere the temperature and the pressure are high(high-temperature/high-pressure period) continues to some extent, thechemical reaction between fuel and oxygen starts. Subsequently thechemical reaction will progress in a chain-reaction manner, andcombustion of the air-fuel mixture will be completed.

FIG. 3 schematically shows a typical gasoline engine comprising a pistonand a cylinder. In a reciprocating gasoline engine as shown in FIG. 3,the inventors studied a condition of an in-cylinder temperature T and anin-cylinder pressure P for causing an air-fuel mixture to self-ignite atan optimal ignition timing (Minimum Advance for Best Torque; hereinafterreferred to as “MBT ignition timing”), while taking into account theaforementioned characteristics of the compression self-ignitioncombustion. Although the MBT ignition timing (optimal ignition timing)varies depending on engine load, it is roughly in the range from atiming corresponding to a compression top dead center (compression TDC)(during high engine loads) to 3 degrees CA ATDC (during low engineloads). As is well known, the “compression TDC” means a top dead centerbetween a compression stroke and a subsequent expansion stroke, and the“degrees CA ATDC” means a crank angle after a top dead center.

As a prerequisite to causing an air-fuel mixture to self-ignite at theMBT ignition timing, it is necessary to control the in-cylindertemperature T and the in-cylinder pressure P just before the MBTignition timing, according to engine speed. Specifically, thehigh-temperature/high-pressure period of the air-fuel mixture becomesshorter as the engine speed becomes higher, and becomes longer as theengine speed becomes lower. Thus, it is necessary to raise thein-cylinder temperature and pressure T, P just before the MBT ignitiontiming, in a high engine speed range where thehigh-temperature/high-pressure period is relatively short, and to lowerthe in-cylinder temperature and pressure T, P just before the MBTignition timing, in a low engine speed region where thehigh-temperature/high-pressure period is relatively long. As describedabove, the MBT ignition timing is limited to a vicinity of a compressionTDC. Thus, in the following description, an in-cylinder temperature andan in-cylinder pressure at a compression TDC will be used as respectiverepresentative examples of the in-cylinder temperature T and thein-cylinder pressure P just before the MBT ignition timing (i.e., justbefore the air-fuel mixture self-ignites), and will be referred torespectively as “compression end temperature Tx” and “compression endpressure Px”.

FIG. 4 is a graph prepared by calculating a condition of the compressionend temperature Tx and the compression end pressure Px (the in-cylindertemperature and the in-cylinder pressure at a compression TDC) forcausing an air-fuel mixture to self-ignite at the MBT ignition timing(optimal ignition timing) when the engine speed is 1000 rpm, based on anelementary reaction calculation and a state equation. In the calculationfor this graph, the excess air ratio λ, i.e., a ratio of an actualair-fuel ratio to the stoichiometric air-fuel ratio, was set to 2.4, anda compression ratio was set to 18. The excess air ratio λ was set to2.4, because the combustion-generated NOx amount (raw NOx amount) itselfcan be sufficiently reduced so as to meet the emission regulations evenif a NOx catalyst is omitted (see the section (1-1)). Further, thecompression ratio was set to 18, because it is necessary to raise atemperature and a pressure of the air-fuel mixture based on acompression ratio greater that that in a standard reciprocating gasolineengine, in order to cause the air-fuel mixture to self-ignite under theultra-lean air-fuel ratio (λ=2.4), and the high-compression ratio isalso advantageous in terms of engine thermal efficiency.

In the graph of FIG. 4, the line L1 is formed by connecting respectivevalues of the compression end temperature Tx and the compression endpressure Px required for self-ignition of the air-fuel mixture at theMBT ignition timing. Thus, it is meant that, if the values Tx, Px areplotted on the line L1, a timing of the self-ignition is coincident withthe MBT ignition timing.

Further, in the graph of FIG. 4, each of the lines M1, M2, M3 representsa value of the engine load based on an amount of fresh air to beintroduced into a cylinder of the engine, wherein the line M1, the lineM2 and the line M3 represent a full engine load, a ⅓ engine load and ano engine load, respectively. Specifically, on the assumption that thecompression ratio is fixed at 18, a combustion chamber volume at acompression TDC is maintained constant. Thus, according to the stateequation, the fresh air amount (engine load) is proportional to thecompression end pressure Px, and inversely proportional to thecompression end temperature Tx. Therefore, the plurality of lines M1 toM3 each having a different slope can be defined for each value of theengine load. In FIG. 4, when each value of the engine load is expressedin indicated mean effective pressure (IMEP), an IMEP at the full engineload (M1), an IMEP at the ⅓ engine load (M2) and an IMEP at the noengine load (M3) are 1300 kPa, 500 kPa and 200 kPa, respectively.

As shown in the graph of FIG. 4, the condition for causing the air-fuelmixture to self-ignite at the MBT ignition timing (line L1) is morelargely shifted to a high-temperature/low-pressure side as the engineload becomes lower, and conversely more largely shifted to alow-temperature/high-pressure side as the engine load becomes higher.Specifically, when the engine load is low and thereby the fresh airamount is small, the molecular collision frequency becomes lower. Thus,in order to induce the chemical reaction, it is necessary to raise thecompression end temperature Tx to increase the molecular velocity. Onthe other hand, when the engine load is high and thereby the fresh airamount is large, the molecular collision frequency becomes higher. Thus,in order to induce the self-ignition at the same timing, it is necessaryto lower the compression end temperature Tx to reduce the molecularvelocity.

If the condition of the compression end temperature Tx and thecompression end pressure Px is deviated to a higher temperature/higherpressure side (upper right side of the graph) than the line L1, theself-ignition timing becomes earlier than the MBT ignition timing. Ifthe temperature-pressure condition are deviated to a lowertemperature/lower pressure side (lower left side of the graph) than theline L1, the self-ignition timing becomes later than the MBT ignitiontiming. Thus, if the temperature-pressure condition is largely deviatedto the higher temperature/higher pressure side or the lowertemperature/lower pressure side with respect to the line L1, preignitionor knocking will occur on the higher temperature/higher pressure side,or misfire will occur on the lower temperature/lower pressure side.

FIG. 5 shows a result obtained by performing the same calculation asthat for FIG. 4 while variously changing only the engine speed on thebasis of the graph of FIG. 4. In the lines L1 to L6 illustrated in FIG.5, the line L1 represents the temperature-pressure condition when theengine speed is 1000 rpm, as with the line L1 in FIG. 4, and theremaining lines L2, L3, L4, L5, L6 represent temperature-pressureconditions when the engine speed is 2000 rpm, 3000 rpm, 4000 rpm, 5000rpm and 6000 rpm, respectively.

As is clear from the graph of FIG. 5, each of the lines L2 to L6representing the temperature-pressure conditions at the respectiveengine speeds 2000 rpm to 6000 rpm slopes downwardly and rightwardly,with a tendency similar to the line L1 at 1000 rpm. Further, the linesL1 to L6 are located in order of L1, L2, - - - , L6 in a rightwarddirection, which shows that it is necessary to raise the compression endtemperature Tx along with an increase in the engine speed, so as toensure the self-ignition. This is because the period where the air-fuelmixture is exposed to high temperatures and high pressures(high-temperature/high-pressure period) becomes shorter as the enginespeed becomes higher, and thereby it is necessary to induce the chemicalreaction within a shorter period of time.

In FIG. 5, each of the lines L1 to L6 has the downward and rightwardsloping, which shows that the temperature-pressure condition for causingthe air-fuel mixture to self-ignite at the MBT ignition timing is morelargely shifted to a high-temperature/low-pressure side as the engineload becomes lower, and more largely shifted to alow-temperature/high-pressure side as the engine load becomes higher,irrespective of engine speed values. However, in reality, if an initialtemperature on a compression stroke (compression initial temperature)and the compression ratio are maintained constant, the compression endtemperature Tx is not changed according to an increase/decrease in theengine load (i.e., increase/reduction in the fresh air amount), althoughonly the compression end pressure Px is changed. Therefore, it issubstantially impossible to create an environment meeting thetemperature-pressure conditions as indicated by the lines L1 to L6 inFIG. 5 (condition in the gray region S), under a constant compressionratio, for example, unless the engine has a device capable of freelyheating and cooling fresh air.

For example, the zone W indicated by the hatched lines in FIG. 5 denotesa range of the compression end temperature Tx and the compression endpressure Px to be obtained by compressing a certain amount of fresh airhaving a normal temperature (which corresponds to a compression initialtemperature of 75° C.) and an atmospheric pressure (0.1 MPa), at acompression ratio of 16 to 18. As seen in FIG. 5, when fresh air havinga normal temperature and an atmospheric pressure is compressed at acompression ratio of 16 to 18, the range denoted by the zone W canpartly meet the temperature-pressure condition for causing the air-fuelmixture to adequately self-ignite (condition in the gray region S).However, for example, if the fresh air amount (engine load) is reducedfrom that in the zone W, the range of the compression end temperature Txand the compression end pressure Px is shifted to a lower pressure sidethan a position of the zone W (to a side just below the zone W), anddeviated from the temperature-pressure condition for causing theair-fuel mixture to adequately self-ignite (region S). Thus, it isproven that the compression self-ignition combustion cannot beadequately produced over the entire engine load/engine speed rangesimply by increasing/reducing the fresh air amount.

If fresh air can be freely heated and cooled, the compressionself-ignition combustion can be adequately produced over the entireengine load/engine speed range in theory. However, as a prerequisite tomeeting the entire temperature-pressure condition as indicated by theregion S in FIG. 5, it is necessary to heat and cool fresh air over afairly-wide temperature range. Thus, such a technique would not berealistic in view of problems in terms of cost, controllability, etc.

(1-3) Solution

Then, the inventors came up with an idea of creating substantially thesame condition as the region S indicative of the condition for theadequate compression self-ignition combustion, by additionallyperforming supercharging and/or internal exhaust gas recirculation(internal EGR) while changing the engine compression ratio in a rangeequal to or less than 18.

As means for changing the compression ratio in the range equal to orless than 18, a technique may be employed which is designed to retard anintake-valve closing timing under a condition that a geometriccompression ratio of an engine is set to 18. When the intake-valveclosing timing is retarded, a compression start timing is retarded, sothat a substantial compression ratio (effective compression ratio) ofthe engine becomes less than 18.

An operation of reducing the effective compression ratio leads tolowering of the compression end temperature Tx, so that it is effectiveparticularly in a low engine speed/high engine load range (upper leftportion of the region S in FIG. 5). However, when the intake-valveclosing timing is retarded so as to reduce the effective compressionratio, the fresh air amount (amount of fresh air to be introduced intothe cylinder) is reduced. Thus, it is anticipated that, if the effectivecompression ratio is simply reduced, the compression end pressure Px isexcessively lowered as compared with the temperature-pressure conditionfor the self-ignition in the low engine speed/high engine load range,which causes insufficiency in torque, or misfire. Therefore, inconjunction with a reduction in the effective compression ratio,supercharging is performed to compensate for a reduction of the freshair amount.

For example, in a high engine speed range (right portion of the regionS) or a low engine load range (lower portion of the region S), it isnecessary to raise the compression end temperature Tx. In this case,internal EGR for causing high-temperature burned gas to remain in thecylinder is performed. Through the internal EGR, it becomes possible toraise the compression end temperature Tx while raising the initialtemperature on a compression stroke (compression initial temperature),so as to create a temperature-pressure condition suitable for thecompression self-ignition combustion in the high engine speed/low engineload region. As means for performing the internal EGR, a technique maybe employed which is designed to provide a negative overlap period whereboth of an intake valve and an exhaust valve are closed during atransition from an exhaust stroke to a subsequent intake stroke.

Embodiments of the Present Invention

(2-1) Overall Configuration

FIG. 6 is a schematic diagram showing an overall configuration of anengine according to one embodiment of the present invention made basedon the above basic theory. The engine illustrated in FIG. 6 is amulti-cylinder gasoline engine including an engine body 1 whichcomprises a cylinder block 3 having a plurality of cylinders 2 (only oneof them is illustrated in FIG. 6) arranged in a direction perpendicularto the drawing sheet, and a cylinder head 4 disposed on the cylinderblock 3. In this embodiment, fuel to be supplied to the engine body 1may be any type consisting mainly of gasoline, wherein a content of thefuel may consist only of gasoline, or may contain gasoline and othercomponent such as ethanol (ethyl alcohol).

A piston 5 is inserted in each of the cylinders 2 of the engine body 1in a reciprocatingly slidable manner. The piston 5 is connected to acrankshaft 7 through a connecting rod 8 to allow the crankshaft 7 to berotated about a central axis thereof according to a reciprocatingmovement of the piston 5.

A combustion chamber 6 is defined just above the piston 5, and thecylinder head 4 is formed with an intake port 9 and an exhaust port 10each opened to the combustion chamber 6, and provided with an intakevalve 11 and an exhaust valve 12 for opening and closing respective onesof the ports 9, 10. Each of the intake valve 11 and the exhaust valve 12is adapted to be openably/closably driven by a respective one of a pairof valve operating mechanisms 13 each including a camshaft (not shown)provided in the cylinder head 4, in conjunction with the rotation of thecrankshaft 7.

A VVL mechanism 14 and a VVT mechanism 15 are incorporated in each ofthe valve operating mechanisms 13 for the intake valve 11 and theexhaust valve 12. The VVL is an abbreviation for Variable Valve Lift,and the VVL mechanism 14 is adapted to variably set a lift amount (valveopening amount) of each of the intake and exhaust valves 11, 12. The VVTis an abbreviation for Variable Valve Timing, and the VVT mechanism 15is adapted to variably set opening and closing timings (phase angles) ofeach of the intake and exhaust valves 11, 12. As for each of the VVLmechanism 14 and the VVT mechanism 15, various types have already beenput into practical use, and commonly known, and detailed descriptionthereof will be omitted here. For example, a type disclosed in JP2007-85241A may be employed.

A spark plug 16 is provided in the cylinder head 4 of the engine body 1in such a manner that it is exposed to the combustion chamber 6 of eachof the cylinders 2 from thereabove. The spark plug 16 is electricallyconnected to an ignition circuit 17 provided on the cylinder head 4, andadapted to generate a spark discharge in response to a supply ofelectric power from the ignition circuit 17 thereto. In the engineaccording to this embodiment, the compression self-ignition combustionis performed throughout the entire engine operating region, and aspark-ignition combustion is basically not performed, as describedlater. However, for example, during engine starting and during extremelycold conditions, the spark-ignition combustion is necessary to eliminatea risk of misfire which is likely to occur in the compressionself-ignition combustion. The spark plug 16 is used at least in such acase.

An injector 18 is provided in the cylinder head 4 in such a manner thatit is exposed to the combustion chamber 6 laterally from an intake sideof the combustion chamber 6. The injector 10 is adapted to inject fuel(consisting mainly of gasoline) therefrom into the combustion chamber 6in an intake stroke, etc., of the engine, so as to mix the injected fuelwith air to form an air-fuel mixture having a desired air-fuel ratio inthe combustion chamber 6.

In the above engine body 1, a geometric compression ratio which isdetermined by a stroke volume (swept volume of the piston 5) and avolume of the combustion chamber at a timing when the piston 5 is in aTDC position, is set to 18.

An intake passage 20 and an exhaust passage 21 are connected torespective ones of the intake port 9 and the exhaust port 10 of theengine body 1. Specifically, the intake passage 20 is adapted to supplycombustion air (fresh air) to the combustion chamber 6 therethrough, andthe exhaust passage 21 is adapted to discharge burned gas (exhaust gas)produced in the combustion chamber 6, to outside the engine body 1therethrough.

The intake passage 20 is provided with a throttle valve 22. In theengine according to this embodiment, the lift amount and the opening andclosing timings of each of the intake and exhaust valves 11, 12 arevariably set by operations of the VVL mechanism 14 and the VVT mechanism15, as described above. Thus, an amount of intake air to be charged intothe combustion chamber 6 can be adjusted to control an engine poweroutput, without selectively opening and closing the throttle valve 22.Therefore, the throttle valve 22 is operated to shut off the intakepassage 20, for example, during emergency stop of the engine, andbasically maintained at a fully-opened position, irrespective of engineoperating state. Based on allowing the throttle valve 22 to bemaintained at the fully-opened position, it becomes possible to achievea reduction in pumping loss.

Fresh air passing through the intake passage 20 is compressed by asupercharger 25 and then supplied to the combustion chamber 6.

The supercharger 25 comprises a compressor 26 disposed inside the intakepassage 20, a turbine 27 disposed inside the exhaust passage 21, acoupling shaft 28 coupling the compressor 26 and the turbine 27together, and an electric motor 29 adapted to rotationally drive thecoupling shaft 28. When the turbine 27 is rotated by receiving exhaustgas energy, the compressor 26 is rotated at a high speed interlockinglywith the turbine 27, so that fresh air passing through the intakepassage 20 is compressed and forcedly supplied to the combustion chamber6. Further, according to need, the electric motor 29 is driven to assistthe rotation of the compressor 26.

The compressor 26 is composed of a relatively-large impeller excellentin compression performance. The supercharger 25 adapted to compressintake air using the large-size compressor 26 can bring out highsupercharging performance, particularly, in a high engine load rangehaving large exhaust gas energy. Further, according to need, therotation assist is performed by the electric motor 29, to allow intakeair to be compressed with excellent response.

The intake passage 20 has a water-cooled intercooler 30 provideddownstream of the compressor 26 and adapted to cool fresh air having atemperature raised by the supercharging.

The exhaust passage 21 is provided with a bypass passage 33 for allowingexhaust gas to bypass the turbine 27, and an electrically-operatedwastegate valve 34 adapted to selectively open and close the bypasspassage 33. Specifically, the wastegate valve 34 is operable toselectively open and close the bypass passage 33 so as to switch betweena first state in which exhaust gas flows through the turbine 27 torotationally drive the turbine 27, and a second state in which exhaustgas bypasses the turbine 27 to stop the rotation of the turbine 27.

The exhaust passage is also provided with a catalytic converter 32 forpurifying exhaust gas. The catalytic converter 32 is provided with athree-way catalyst housed therein, and adapted to purify harmfulcomponents contained in exhaust gas passing therethrough by an action ofthe three-way catalyst.

(2-2) Control System

The above engine further comprises an ECU (Engine Control Unit) 40composed, for example, of a conventional CPU, a conventional memory andothers, to serve as control means (controller) for comprehensivelycontrolling an operation of the engine.

The ECU 40 is electrically connected to a plurality of sensors installedat respective positions of the engine. More specifically, the ECU 40 iselectrically connected to each of an engine speed sensor 51 fordetecting a rotation speed of the crankshaft 7, an airflow sensor 52 fordetecting an amount of fresh air passing through the intake passage 20,an accelerator pedal angle sensor 53 for detecting a depression amount(depression angle) of an accelerator (not shown) adapted to be depressedby a driver. Each of the sensors 51 to 53A is operable to input adetection value into the ECU 40 in the form of an electric signal.

Further, the ECU 40 is electrically connected to each of the VVLmechanism 14, the VVT mechanism 15, the ignition circuit 17, theinjector 18, the throttle valve 22, the electric motor 29, and thewastegate valve 34, and adapted to output a driving control signal toeach of these components.

Specific functions of the ECU 40 will be described below. As majorfunctional elements, the ECU 40 has valve control means 41,supercharging control means 42 and injector control means 43.

The valve control means 41 is designed to drive the VVL mechanism 14 andthe VVT mechanism 15 to variably set a lift characteristic (opening andclosing timings and a lift amount) of each of the intake and exhaustvalues 11, 12. More specifically, based on changing the liftcharacteristic of each of the intake and exhaust values 11, 12, thevalve control means 41 has a function of controlling an amount of burnedgas remaining in the combustion chamber 6 (internal EGR amount), and afunction of controlling an effective compression ratio of the engine.

In an operation of controlling the internal EGR amount, the valvecontrol means 41 changes the lift characteristic of each of the intakeand exhaust values 11, 12, for example, in a manner as shown in FIG. 7.In FIG. 7, the lines Lex represent the lift characteristic of theexhaust value 12, and the lines Lin represent the lift characteristic ofthe intake value 11. Further, the range NVO on the horizontal axisrepresents a negative overlap period where both of the intake andexhaust valves 11, 12 are closed during transition from an exhauststroke to a subsequent intake stroke. The valve control means 41 isoperable to drive the VVL mechanism 14 and the VVT mechanism 15 tochange the lift characteristic of each of the intake and exhaust values11, 12 so as to increase/reduce the negative overlap period NVO toadjust internal EGR amount (amount of burned gas remaining in thecombustion chamber 6).

In an operation of controlling the effective compression ratio of theengine, the valve control means 41 changes the lift characteristic ofthe intake valve 11, for example, in a mode as shown in FIG. 8.Specifically, the intake valve 11 is normally closed at a timingadjacent to and on a retard side of an intake BDC (bottom dead center ofan intake stroke) (a timing just after an intake BDC), as indicated bythe solid waveform in FIG. 8. In this state, the effective compressionratio is identical to the geometric compression ratio (in thisembodiment, 18). When the closing timing of the intake valve 11 is setto a largely retarded point with respect to the intake BDC, as indicatedby the broken line in FIG. 8, a start timing of a subsequent compressionstroke is retarded, and thereby the effective compression ratio(substantial compression ratio) of the engine is reduced. The valvecontrol means 41 is operable to increase/reduce an amount of retard(retard amount) of the closing timing of the intake valve 11 so as tovariably set the effective compression ratio of the engine.

As above, FIG. 8 illustrate one example in which the VVT mechanism 15for the intake valve 11 is activated to retard an operation timing(opening and closing timings) of the intake valve 11 to a retard side asindicated by the broken line in FIG. 8. When only the VVT mechanism 15is activated in the above manner, not only the closing timing but alsothe opening timing of the intake valve 11 are changed, and thereby thenegative overlap period NOV illustrated in FIG. 7 is simultaneouslychanged. Thus, in cases where it is desirable to change only theeffective compression ratio without changing the negative overlap periodNOV, both of the VVL mechanism 14 and the VVT mechanism 15 may beactivated to control the lift amount and the opening and closing timingsin such a manner that only the closing timing of the intake valve 11 isvariably changed while maintaining the opening timing of the intakevalve 11 constant.

The supercharging control means 42 is designed to control thesupercharger 25 by driving the electric motor 29 for the supercharger 25according to need, and openably/closably driving the wastegate valve 34,so as to obtain an adequate supercharging pressure.

The injector control means 43 is designed to control an injection timingand an injection amount (injection period) of fuel to be injected fromthe injector 18 into the combustion chamber 6. More specifically, theinjector control means 43 has a function of controlling an air-fuelratio in the cylinder by calculating a target fuel injection amount forobtaining a given air-fuel ratio, based on information, such as anintake air amount (fresh air amount) input from the airflow sensor 52,and opening the injector 18 only for a time corresponding to the targetfuel injection amount. In this embodiment, the injector control means 43is operable to control the fuel injection amount from the injector 18 toallow an excess air ratio λ which is a ratio of an actual air-fuel ratioto the stoichiometric air-fuel ratio to be maintained at 2.4 throughoutthe entire engine operating region. As for the fuel injection timing,fuel is injected during an intake stroke to sufficiently ensure a timefor mixing between fuel and air.

(2-3) Specific Example of Control

How to control the engine depending engine load and engine speed by theabove ECU 40 will be specifically described below. On an assumption thatan engine operating state in which the engine speed is 1000 rpm and theengine load is a ⅓ engine load (IMEP=500 kPa), i.e., the point R0 inFIG. 5, is defined as a representative point, the following descriptionwill be made about a specific example of a control scheme to be executedwhen the engine load or engine speed is changed from the representativepoint R0.

FIG. 9 is an engine operating zone map, wherein the horizontal axisrepresents the engine speed, and the vertical axis represents the engineload (IMEP). Each of the arrowed lines A1, A2 in FIG. 9 denotes that theengine operating state is changed along a direction of the engine loadaxis (engine load-axis direction), and FIG. 10 shows how each of theeffective compression ratio ε′, an initial temperature T0 on acompression stroke (compression initial temperature T0) and an initialpressure P0 on the compression stroke (compression initial pressure P0)is changed according to the change in engine load along the engineload-axis direction. Further, the arrowed line A3 in FIG. 9 denotes thatthe engine operating state is changed along a direction of the enginespeed axis (engine speed-axis direction), and FIG. 11 shows how each ofthe effective compression ratio ε′, the compression initial temperatureT0 and the compression initial pressure P0 is changed according to thechange in engine speed along the engine speed-axis direction. The terms“compression initial temperature T0” and “compression initial pressureP0” here are defined, respectively, as an in-cylinder temperature and anin-cylinder pressure at a timing when the intake valve 11 is closed.Further, in cases where the internal EGR for causing burned gas toremain in the cylinder (combustion chamber 6) is performed, the term“compression initial pressure P0” is defined as a partial pressure offresh air in the cylinder (a value obtained by subtracting a partialpressure of burned gas from a total pressure in the cylinder) at aclosing timing of the intake valve 11.

Firstly, a control scheme at the representative point R0 will bedescribed. As shown in FIGS. 10 and 11, in the engine according to thisembodiment, when the engine operating state is at the representativepoint R0 (engine speed=1000 rpm and IMEP=500 kPa), the effectivecompression ratio ε′ is set to 15 which is less than the geometriccompression ratio (=18), and the compression initial pressure P0 is setto be greater than a pressure in a non-supercharging mode (i.e.,non-supercharging pressure; atmospheric pressure of 0.1 MPa) by a givenvalue. On the other hand, the compression initial temperature T0 ismaintained at a normal temperature (75° C.).

More specifically, at least the closing timing of the intake valve 11 isretarded by the valve control means 41, to reduce the effectivecompression ratio ε′ from 18 to 15, and the supercharger 25 is driven bythe supercharging control means 42, to set the compression initialpressure P0 to be greater than the non-supercharging pressure by a givenvalue. The compression initial temperature T0 is maintained at thenormal temperature, as mentioned above. Thus, the internal EGR (controlfor forming the negative overlap period to cause high-temperature burnedgas to remain in the combustion chamber 6) is not performed. As above,the control scheme for reducing the effective compression ratio ε′ andraising the compression initial pressure P0 (i.e., supercharging freshair) is executed, because, in the graph of FIG. 5, a position of therepresentative point R0 is located slightly on a lowertemperature/higher pressure side than the zone W which is atemperature-pressure range to be obtained when fresh air having theatmospheric pressure is compressed at a compression ratio of 16 to 18

In other words, as a prerequisite to causing an air-fuel mixture toself-ignite at the representative point R0 (engine speed=1000 rpm andIMEP=500 kPa) to adequately perform the compression self-ignitioncombustion, it is necessary to allow a condition of the compression endtemperature Tx and the compression end pressure Px to correspond to theposition of the representative point R0 in FIG. 5. However, atemperature-pressure condition corresponding to the representative pointR0 is unable to be obtained simply by compressing fresh air having theatmospheric pressure, at a compression ratio of 16 to 18, as in the zoneW, and it is necessary to shift the condition of the compression endtemperature Tx and the compression end pressure Px slightly to a lowertemperature/higher pressure side. Therefore, as shown in FIGS. 10 and11, at the representative point R0, the effective compression ratio ε′is reduced to 15 while raising the compression initial pressure P0(fresh air amount) by the supercharging, so as to lower the compressionend temperature Tx and raise the compression end pressure Px, ascompared with the zone W. In this manner, the temperature-pressurecondition capable of adequately performing the compression self-ignitioncombustion at the representative point R0 (engine speed=1000 rpm andIMEP=500 kPa) can be created.

Secondly, a control scheme to be executed when the engine operatingstate is changed from the representative point R0 along the engineload-axis direction (see the arrowed lines A1, A2 in FIG. 9) will bedescribed. When the engine operating state is shifted from therepresentative point R0 to a higher engine load side as indicated by thearrowed line A1, a control scheme for gradually reducing the effectivecompression ratio ε′ while gradually raising the compression initialpressure P0, according to an increase in the engine load, is executed,as shown in the range SA1 illustrated in FIG. 10. On the other hand, thecompression initial temperature T0 is maintained at the normaltemperature, irrespective of engine load values.

Specifically, when the engine operating state is shifted from therepresentative point R0 to a higher engine load side, the closing timingof the intake valve 11 is further retarded as compared with that for therepresentative point R0, to gradually reduce the effective compressionratio ε′ in a range less than 15, and finally the effective compressionratio ε′ is reduced down to 10.5 at an operating point R1 corresponding,for example, to the full engine load (IMEP=1300 kPa). Further, alongwith a reduction in the effective compression ratio ε′, thesupercharging pressure based on the supercharger 25 is gradually raisedto further raise the compression initial pressure P0 as compared withthat for the representative point R0. On the other hand, the compressioninitial temperature T0 is maintained at the normal temperature withoutperforming the internal EGR.

As above, on a higher engine load side than the representative point R0,the effective compression ratio ε′ is gradually reduced while raisingthe compression initial pressure P0 by the supercharging (i.e.,increasing the fresh air amount), along with an increase in the engineload. This corresponds to changing the temperature-pressure conditionupwardly along the line L1 on the graph of FIG. 5 (see the arrowed lineA1 a). Thus, based on controlling the temperature/pressure in the abovemanner, the compression self-ignition combustion can be adequatelyperformed in the high engine load range.

Specifically, as a prerequisite to allowing the adequate compressionself-ignition combustion (combustion mode where an air-fuel mixtureself-ignites at the MBT ignition timing) to be performed on a higherengine load side than the representative point R0, it is necessary toshift the condition of the compression end temperature Tx and thecompression end pressure Px to the lower temperature/higher pressureside along with an increase in the engine load, as indicated by thearrowed line A1 a extending upwardly along the line L1. For thispurpose, in FIG. 10, when the engine load is increased from therepresentative point R0, the effective compression ratio ε′ is reducedwhile raising the compression initial pressure P0 (fresh air amount) bythe supercharging. This makes it possible to lower the compression endtemperature Tx while raising the compression end pressure Px, withrespect to the representative point R0, as indicated by the arrowed lineA1 a in FIG. 5, so as to continue the adequate compression self-ignitioncombustion toward a higher engine load side.

Thirdly, a control scheme to be executed when the engine operating stateis changed from the representative point R0 on a lower engine load sideas indicated by the arrowed line A2 in FIG. 9 will be described. In thiscase, a control scheme for gradually increasing the effectivecompression ratio ε′ and raising the compression initial temperature T0while gradually lowering the compression initial pressure P0, accordingto an increase in the engine load, is executed, as shown in the rangeSA2 illustrated in FIG. 10.

Specifically, when the engine operating state is shifted from therepresentative point R0 to a lower engine load side, a retard amount ofthe closing timing of the intake valve 11 is set to be less than thatfor the representative point R0, to gradually increase the effectivecompression ratio ε′ in a range equal to or greater than 15, and finallythe effective compression ratio ε′ is increased up to 18 equal to thegeometric compression ratio, at an operating point R2 corresponding, forexample, to the no engine load (IMEP=200 kPa). Further, in conjunctionwith a reduction in the effective compression ratio ε′, the internal EGRis performed to gradually increase an amount of burned gas remaining inthe cylinder (combustion chamber 6) so as to gradually raise thecompression initial temperature T0 in a range equal to or greater thanthe normal temperature. In addition, the supercharging using thesupercharger 25 is stopped when the engine operating state is shiftedslightly on a lower engine load side than the representative point R0,and the internal EGR is performed in the above manner to lower a partialpressure of fresh air in the cylinder, so that the compression initialpressure P0 is lowered down to less than the atmospheric pressure alongwith a decrease of the engine load.

As above, on a lower engine load side than the representative point R0,the effective compression ratio ε′ is gradually increased while loweringthe compression initial pressure P0 by the internal EGR (i.e., reducingthe fresh air amount), along with a decrease in the engine load. Thiscorresponds to changing the temperature-pressure condition downwardlyalong the line L1 on the graph of FIG. 5 (see the arrowed line A2 a).Thus, based on controlling the temperature/pressure in the above manner,the compression self-ignition combustion can be adequately performed inthe low engine load range.

Specifically, as a prerequisite to allowing the adequate compressionself-ignition combustion (combustion mode where an air-fuel mixtureself-ignites at the MBT ignition timing) to be performed on a lowerengine load side than the representative point R0, it is necessary toshift the condition of the compression end temperature Tx and thecompression end pressure Px to the higher temperature/lower pressureside along with a decrease in the engine load, as indicated by thearrowed line A2 a extending downwardly along the line L1. For thispurpose, in FIG. 10, when the engine load is decreased from therepresentative point R0, the effective compression ratio ε′ is increasedwhile lowering the compression initial pressure P0 (fresh air amount) bythe internal EGR. This makes it possible to raise the compression endtemperature Tx while lowering the compression end pressure Px, withrespect to the representative point R0, as indicated by the arrowed lineA2 a in FIG. 5, so as to continue the adequate compression self-ignitioncombustion toward a lower engine load side.

Fourthly, a control scheme to be executed when the engine operatingstate is changed from the representative point R0 to a higher enginespeed side as indicated by the arrowed line A3 in FIG. 9 will bedescribed. In this case, a control scheme for raising the compressioninitial temperature T0 along with an increase in the engine speed isperformed, as shown in FIG. 11. Specifically, the internal EGR isperformed to gradually increase the burned gas amount (amount of burnedgas remaining in the cylinder) so as to gradually raise the compressioninitial temperature T0 in a range equal to or greater than the normaltemperature.

As for the effective compression ratio ε′ and the compression initialpressure P0, when the engine speed is in the range of 1000 to 2000 rpm,the effective compression ratio ε′ is set to be less than 18 (15 to lessthan 18), and the compression initial pressure P0 is set to be greaterthan the atmospheric pressure, by the supercharging. Further, in therange of 2000 to 6000 rpm which is a higher engine speed side than theabove engine speed range, the effective compression ratio ε′ ismaintained at 18 equal to the geometric compression ratio, and thesupercharging is performed at a relatively low level to set thecompression initial pressure P0 to be slightly greater than theatmospheric pressure.

As above, in a higher engine speed side than the representative pointR0, the compression initial temperature T0 is raised by the internalEGR. This corresponds to changing the temperature-pressure conditionrightwardly along the line M2 on the graph of FIG. 5 (see the arrowedline A3 a). Thus, based on controlling the temperature/pressure in theabove manner, the compression self-ignition combustion can be adequatelyperformed in the high engine speed range.

Specifically, as a prerequisite to allowing the adequate compressionself-ignition combustion (combustion mode where an air-fuel mixtureself-ignites at the MBT ignition timing) to be performed in a higherengine speed side than the representative point R0, it is necessary toprimarily raise the compression end temperature Tx along with anincrease in the engine speed, as indicated by the arrowed line A3 aextending rightwardly along the line M2. For this purpose, in FIG. 11,when the engine speed is increased from the representative point R0, thecompression initial temperature T0 is raised by the internal EGR. Thismakes it possible to raise the compression end temperature Tx withrespect to the representative point R0, as indicated by the arrowed lineA3 a in FIG. 5, so as to continue the adequate compression self-ignitioncombustion toward a higher engine speed side.

However, particularly in a higher engine speed side than 2000 rpm, theinternal EGR amount is increased to raise the compression initialtemperature T0 under the condition that the effective compression ratioε′ is fixed to 18, so that a density of fresh air becomes lower alongwith a rise in the temperature due to the internal EGR, which is likelyto cause reduction in engine power output. For this reason, in theexample illustrated in FIG. 11, in the higher engine speed side than2000 rpm, the supercharging is performed at a relatively low level toslightly raise the compression initial pressure P0, so as to avoidlowering of the fresh air density to adequately ensure the engine poweroutput.

The control schemes in the above examples have been described based onFIGS. 10 and 11, respectively, on the assumption that, with respect tothe representative point R0, only the engine load is changed under aconstant engine speed (as the arrowed line A1 and A2 in FIG. 9), andonly the engine speed is changed under a constant engine load (as thearrowed line A3 in FIG. 9). However, even when the engine operatingstate is shifted to any zone of the engine operating region other thanthose in the above examples, the adequate compression self-ignitioncombustion can be performed throughout the entire engine operatingregion by deriving a condition of the compression end temperature Tx andthe compression end pressure Px suitable for each engine operating zone,from the graph of FIG. 5, and controlling the effective compressionratio ε′, the compression initial temperature T0 and/or the compressioninitial pressure P0 so as to establish the suitable condition.

More specifically, although FIG. 10 illustrates the control scheme to beperformed when only the engine load is changed under a constant enginespeed of 1000 rpm, a control pattern depending on the engine load is thesame at any engine speed other than 1000 rpm. For example, in caseswhere the engine speed is set to a constant value, such as 2000 rpm,3000 rpm, 4000 rpm, 5000 rpm or 6000 rpm, a control scheme may beemployed which is configured to, on a higher engine load side than agiven engine load, reduce the effective compression ratio ε′ whileincreasing the fresh air amount (i.e., raising the compression initialpressure P0) by the supercharging, and, on a lower engine load side thanthe given engine load, increasing the effective compression ratio ε′while reducing the fresh air amount (i.e., lowering the compressioninitial pressure P0) by the internal EGR, as with the control schemeillustrated in FIG. 10.

Further, although FIG. 11 illustrates the control scheme to be performedwhen only the engine speed is changed under a constant engine load(IMEP) of 500 kPa, a control pattern depending on the engine speed isthe same at any engine load (IMEP) other than 500 kPa. Specifically, atany engine load (IMEP) other than 500 kPa, the compression initialtemperature T0 may be raised by the internal EGR while performing thesupercharging according to need, along with an increase in the enginespeed, as with the control scheme illustrated in FIG. 11.

However, in the above cases, specific control target values of theeffective compression ratio ε′, the compression initial temperature T0,etc., have to be set differently from those in the control schemesillustrated in FIGS. 10 and 11. For example, in FIG. 9, the zone Urepresenting a high engine speed/low engine load range corresponds to alower right portion of the region S as cross-referred with FIG. 5. Inthis engine operating zone, it is necessary to maximally raise thecompression end temperature Tx. Thus, the effective compression ratio ε′is set to the maximum value of 18 while maximally increasing theinternal EGR amount to further raise the compression initial temperatureT0. Specifically, as shown in FIGS. 10 and 11, at the operating point R2(engine speed=1000 rpm and no engine load (IMEP=200 kPa)) and theoperating point R3 (engine speed=6000 rpm and ⅓ engine load (IMEP=500kPa)), the internal EGR is performed to raise the compression initialtemperature T0 so as to raise the compression end temperature Tx.Differently, in the high engine speed/low engine load range (zone U inFIG. 9), the internal EGR is performed to obtain a larger internal EGRamount than those for the operating points R2, R2, so as to furtherraise the compression end temperature Tx. This makes it possible toperform the adequate the compression self-ignition combustion in thehigh engine speed/low engine load range.

In FIG. 9, the zone V representing a high engine speed/high engine loadrange corresponds to an upper right portion of the region S ascross-referred with FIG. 5. In this engine operating zone, it isnecessary to set the compression end temperature Tx to a higher value,for example, as compared with the low engine speed/high engine loadoperating point R1 (engine speed=1000 rpm and IMEP=1300 kPa). For thispurpose, in the high engine speed/high engine load range, the effectivecompression ratio ε′ may be set to a larger value than that for theoperating point R1. Specifically, at the operating point R1, in order tocreate the condition that the compression end temperature Tx isrelatively low and the compression end pressure Px is relatively high,the effective compression ratio ε′ is reduced down to 10.5 while raisingthe compression initial pressure P0 by the supercharging. Differently,in the high engine speed/high engine load range (zone V in FIG. 9),considering the need for more largely raising the compression endtemperature Tx as compared with the operating point R1, the effectivecompression ratio ε′ may be set to be greater than 10.5 so as to providea higher in-cylinder temperature.

(2-4) Functions, Effects, etc.

As described above, in the engine according to the above embodiment,wherein the excess air ratio λ, i.e., a ratio of an actual air-fuelratio to the stoichiometric air-fuel ratio, is set to 2.4 throughout theentire engine operating region, as a control scheme to be performed whenthe engine operating state is changed along the engine speed-axisdirection, for example, as shown in FIG. 11, in a higher engine speedside, the compression initial temperature T0 is raised by increasing theinternal EGR amount (amount of burned gas remaining in the cylinder)with respect to the lower engine speed side. This control scheme has anadvantage of being able to allow the compression self-ignitioncombustion under a lean air-fuel ratio to be performed in a wider enginespeed range so as to more effectively enhance the engine thermalefficiency.

Specifically, in the above embodiment, when the engine operating stateis changed in the engine speed-axis direction, the internal EGR amountis increased as the engine speed becomes higher, to increase thecompression initial temperature T0, so that primarily the compressionend temperature Tx can be raised, as indicated by the arrowed line A3 ain FIG. 5, so as to increase a collision velocity between fuel andoxygen molecules to facilitate a chemical reaction therebetween. Thismakes it possible to reliably cause an air-fuel mixture to self-igniteeven in the high engine speed range where a period having highin-cylinder temperatures/pressures (high-temperature/high-pressureperiod) becomes shorter, so that an engine speed range capable ofperforming the compression self-ignition combustion under a leanair-fuel ratio can be expanded to a higher side to further effectivelyenhance the engine thermal efficiency.

Particularly, in the above embodiment, in conjunction with raising thecompression initial temperature T0 by the internal EGR as shown in FIG.11, the compression initial pressure P0 is raised by the supercharging.This provides an advantage of being able to compensate for lowering offresh air density due to the temperature rise, based on superchargedfresh air, so as to adequately ensure the engine power output,irrespective of engine speed values.

In the above embodiment, the geometric compression ratio is set to 18.Further, on a higher engine load side than a given engine load (e.g.,IMEP=500 kPa at an engine speed of 1000 rpm), the fresh air amount isincreased by increasing the supercharging pressure based on thesupercharger 25, while reducing the effective compression ratio ε′, ascompared with a lower engine load side than the given engine load, andon the lower engine load side than the given engine load, the fresh airamount is reduced by the internal EGR while increasing the effectivecompression ratio ε′, as compared with the higher engine load side thanthe given engine load. This control scheme has an advantage of beingable to allow the compression self-ignition combustion under a leanair-fuel ratio to be performed in a wider engine load range so as tomore effectively enhance the engine thermal efficiency.

Specifically, in the above embodiment, the compression end temperatureTx can be lowered while raising the compression end pressure Px, asindicated by the arrowed line A1 a in FIG. 5, by increasing the freshair amount based on the supercharging and reducing the effectivecompression ratio ε′, according to an increase in the engine load, sothat it becomes possible to sufficiently ensure the engine power outputbased on a large amount of supercharged fresh air, while reducing thecollision velocity between fuel and oxygen molecules to suppress thechemical reaction therebetween so as to effectively prevent abnormalcombustion, such as knocking or preignition.

Further, when the engine load is decreased, the compression endtemperature Tx can be raised while lowering the compression end pressurePx, as indicated by the arrowed line A2 a in FIG. 5, by reducing thefresh air amount based on the internal EGR and increasing the effectivecompression ratio ε′, so that it becomes possible to increase amolecular velocity to facilitate the chemical reaction so as to reliablycause an air-fuel mixture to self-ignite, even in a situation where thefresh air amount is small and thereby a frequency of the collisionbetween fuel and the oxygen molecules is low.

Based on the above advantageous effects, in the above embodiment, thecompression self-ignition combustion under a lean air-fuel ratio can beadequately performed in a wider range along the engine load-axisdirection to effectively enhance the engine thermal efficiency.

Particularly in the above embodiment, the internal EGR for causinghigh-temperature burned gas to remain in the combustion chamber 6 isperformed to reduce the fresh air amount on the lower engine load sidethan the given engine load, so that it becomes possible to not onlyreduce the fresh air amount but also raise the compression initialtemperature T0, by internal EGR. This provides an advantage of beingable to effectively create a condition of the compression endtemperature Tx and the compression end pressure Px suitable for a lowengine load so as to allow the compression self-ignition combustionunder a lean air-fuel ratio to be adequately performed on the lowerengine load side.

As above, in the engine according to the above embodiment, respectiveparameters of the effective compression ratio ε′, the compressioninitial temperature T0 and the compression initial pressure P0 can beadequately controlled by taking advantage of supercharging, internalEGR, etc., to create the condition of the compression end temperature Txand the compression end pressure Px as shown in FIG. 5 so as to allowthe adequate compression self-ignition combustion (combustion mode wherean air-fuel mixture self-ignites at the MBT ignition timing) to beperformed throughout the entire engine operating region.

In a verification test using actual engines, the inventors have verifiedthat it is able to cause an air-fuel mixture to self-ignite around theMBT ignition timing throughout the entire engine operating region bycontrolling an engine under the conditions based on the aboveembodiment. Further, in experimental tests, a combustion period of thecompression self-ignition combustion has also be checked, and it hasbeen verified that a so-called “10-to-90% mass burning period (a periodafter 10% of a mass of fuel is burned through until 90% of the mass isburned) falls within a combustion period of about 10 degrees CA. Thiscombustion mode where the 10-to-90% mass burning period falls within 10degrees CA is shorter in combustion period as compared with theconventional spark-ignition combustion mode, and thereby it is expectedto provide higher engine thermal efficiency.

(2-5) Other Embodiments

In the above embodiment, in order to reliably cause an air-fuel mixtureto self-ignite even in a situation where the engine speed is increasedup to the high engine speed range (i.e., thehigh-temperature/high-pressure period is short), the compression endtemperature Tx just before ignition is raised by increasing a burned-gasremaining amount based on the internal EGR (internal EGR amount) toraise the compression initial temperature T0 in the cylinder, forexample, as shown in FIG. 11. However, the technique for rising thecompression end temperature Tx is not limited to the type based on theinternal EGR. For example, it is contemplated to raise the compressioninitial temperature T0 and the compression end temperature Tx byforcedly heating fresh air in the intake passage and introducing theheated air into the cylinder.

Alternatively, the compression end temperature Tx may be raised byactivating the spark plug in a compression stroke to secondarily ignitea part of an air-fuel mixture. Based on raising the compression endtemperature Tx by means of the secondary ignition in this manner, theremaining air-fuel mixture is allowed to self-ignite around thecompression TDC (MBT). In this case, only the compression endtemperature Tx is raised without raising the compression initialtemperature T0.

The control scheme in the above embodiment is configured to, when theengine load is increased to some extent (e.g., in the range equal to orgreater than IMEP=500 kPa as shown in FIG. 10), lower the compressionend temperature Tx while raising the compression end pressure Px, asindicated by the arrowed line A1 a in FIG. 5, by increasing thesupercharging pressure based on the supercharger 25 while graduallyreducing the effective compression ratio ε′. In this control scheme,particularly, when the engine load is increased to a vicinity of thefull engine load (IMEP=1300 kPa), it becomes necessary to introduce alarge amount of fresh air into the cylinder by the supercharging whilelargely reducing the effective compression ratio ε′ with respect to thegeometric compression ratio (e.g., reducing from 18 to 10.5 in the fullengine load). If the supercharger 25 has a sufficient capacity, suchcontrol can be performed without any problem. However, in cased wherethe capacity of the supercharger 25 cannot be sufficiently ensured dueto restrictions of cost, etc., it is assumed that it becomes difficultto achieve such control in the high engine load range. Thus, in a highload range, the combustion mode may be switched to the spark-ignition(SI) combustion mode where an air-fuel mixture is forcedly ignited by aspark discharge from the spark plug 16, wherein the compressionself-ignition combustion is performed only in a partial engine loadrange.

In the above embodiment, the excess air ratio λ, i.e., a ratio of anactual air-fuel ratio to the stoichiometric air-fuel ratio, is set to2.4 throughout the entire engine operating region without exception tosufficiently reduce an amount of NOx to be produced from combustion (rawNOx amount) itself so as to sufficiently meet emission regulations evenif a NOx catalyst is omitted. However, in cases where a NOx catalyst canbe provided in the engine, the excess air ratio λ may be set to be lessthan 2.4 at lest in a part of the engine operating region. However, inview of sufficiently meeting strict emission regulations anticipated inthe future, even if a NOx catalyst is provided, the excess air ratio λshould be set to 2 or more. As described based on FIG. 1, as long as theexcess air ratio λ≧2, NOx emissions can be reduced to a sufficient levelby purifying the produced NOx through the NOx catalyst.

In the above embodiment, the geometric compression ratio of the engineis set to 18. Alternatively, it may be set to any suitable value otherthan 18. For example, in the graph of FIG. 5, in view of the fast thatthe zone W representing a temperature-pressure range to be obtained whenfresh air having the normal temperature and the atmospheric pressure iscompressed at a compression ratio of 16 to 18, overlaps the region Srepresenting the temperature-pressure condition required for theadequate compression self-ignition combustion, it is believed that, aslong as the geometric compression ratio is set to at least 16 or more,the adequate compression self-ignition combustion can be performed ineach engine operating zone as with the above embodiment, byappropriately changing the effective compression ratio ε′ under acondition that the compression ratio 16 is defined as a maximum value ofa control range. For this reason, the geometric compression ratio may beset to at least 16 or more.

In the above embodiment, in order to raise the compression endtemperature Tx to cause an air-fuel mixture to self-ignite, when theengine load is low or the engine speed is high, the negative overlapperiod NVO (FIG. 7) where both of the intake and exhaust valves areclosed during a transition from an exhaust stroke to a subsequent intakestroke, and the internal EGR for causing high-temperature burned gas toremain in the combustion chamber 6 is performed by means of the negativeoverlap period NVO. However, a technique for performing the internal EGRis not limited to the type utilizing the negative overlap period NVO.For example, in an engine having two exhaust valves 12 per cylinder, atechnique may be employed which is designed to open one of the exhaustvalves during an intake stroke to allow burned gas to return from anexhaust passage to a combustion chamber 6 on the intake stroke so as tocause the burned gas to remain in the combustion chamber 6.

At the end of description, features and advantages of the presentinventions disclosed based on the above embodiments will be summarized.

As one aspect of the present invention to achieve the aforementionedobject, there is provided a method for controlling an engine. The methodcomprises a step of allowing a compression self-ignition combustionunder an air-fuel ratio leaner than a stoichiometric air-fuel ratio tobe performed at least in a partial-load range of the engine, wherein,under a condition that an engine speed varies at a same load in anengine operating region of the compression self-ignition combustion, acompression end temperature, which is an in-cylinder temperature justbefore an air-fuel mixture self-ignites, is controlled to be raisedhigher in a higher engine speed side than in a lower engine speed side.

As another aspect of the present invention, there is provided anapparatus for controlling an engine. The apparatus comprises acontroller adapted to control respective sections of the engine to allowa compression self-ignition combustion under an air-fuel ratio leanerthan a stoichiometric air-fuel ratio to be performed at least in apartial-load range of the engine, wherein the controller is operable,under a condition that an engine speed varies at a same load in anengine operating region of the compression self-ignition combustion, tocontrol a compression end temperature, which is an in-cylindertemperature just before an air-fuel mixture self-ignites, in such amanner that it is raised higher in a higher engine speed side than in alower engine speed side.

In the method and apparatus of the present invention, in order to allowthe compression self-ignition combustion under an air-fuel ratio leanerthan the stoichiometric air-fuel ratio to be performed, the compressionend temperature is raised higher in the higher engine speed side than inthe lower engine speed side, so that it becomes possible to increase acollision velocity between fuel and oxygen molecules as the engine speedbecomes higher and thereby the high-temperature/high-pressure period inthe cylinder becomes shorter, to facilitate a chemical reactiontherebetween so as to reliably cause an air-fuel mixture toself-ignition. This makes it possible to expand an engine speed rangecapable of performing the compression self-ignition combustion under alean air-fuel ratio, to a higher engine speed side to furthereffectively enhance engine thermal efficiency.

Preferably, in the method of the present invention, the control forraising compression end temperature includes a control for raising acompression initial temperature which is an in-cylinder temperature at astart timing of a compression stroke, wherein, the compression initialtemperature is controlled to be raised higher in the higher engine speedside than in the lower engine speed side.

Preferably, in the apparatus of the present invention, the control forraising the compression end temperature includes a control for raising acompression initial temperature which is an in-cylinder temperature at astart timing of a compression stroke, wherein the controller is operableto control the compression initial temperature in such a manner that itis raised higher in the higher engine speed side than in the lowerengine speed side.

In the above method and apparatus, the compression end temperature israised by raising the compression initial temperature, so as to allowthe compression self-ignition combustion to be adequately performed inthe high engine speed side.

Preferably, in the above method, the control for raising the compressioninitial temperature includes a control for performing internal EGR forcausing burned gas to remain in a cylinder, wherein, an internal EGRamount is controlled to be increased larger in the higher engine speedside than in the lower engine speed side.

Preferably, in the above apparatus, the control for raising thecompression initial temperature includes a control for performinginternal EGR for causing burned gas to remain in a cylinder, wherein thecontroller is operable to control an internal EGR amount in such amanner that it is increased larger in the higher engine speed side thanin the lower engine speed side.

In the above method and apparatus, the compression initial temperaturecan be adequately raised to a high value by allowing high-temperatureburned gas to remain in the cylinder.

Preferably, in the above method, when the internal EGR is performed at asame engine speed, the internal EGR amount is increased as the loadbecomes lower.

Preferably, in the above apparatus, the controller is operable, when theinternal EGR is performed at a same engine speed, to increase theinternal EGR amount as the load becomes lower.

In the above method and apparatus, even in a situation where the engineload is low (i.e., the fresh air amount is small) and thereby afrequency of the collision between fuel and the oxygen molecules is low,a molecular velocity can be increased to promote a chemical reactiontherebetween so as to reliably cause an air-fuel mixture to self-ignite.

Preferably, the above method comprises a step of raising a compressioninitial pressure which is an in-cylinder pressure at a start timing of acompression stroke, according to a rise in the compression initialtemperature, by means of supercharging.

Preferably, in the above apparatus, the engine is equipped with asupercharger operable to supercharge fresh air, wherein the controlleris operable to activate the supercharger to raise a compression initialpressure which is an in-cylinder pressure at a start timing of acompression stroke, according to a rise in the compression initialtemperature.

In the above method and apparatus, an engine power output can beadequately ensured irrespective of engine speed values by compensatingfor lowering of fresh air density due to the rise in the compressioninitial temperature, based on the supercharging.

Preferably, in the method and apparatus of the present invention, ageometric compression ratio of the engine is 16 or more.

In the above method and apparatus, a geometric compression ratio of theengine is set to 16 or more, so that an air-fuel mixture can besufficiently increased in temperature/pressure so as to reliably causethe air-fuel mixture to self-ignite. In addition, such a highcompression ratio is also advantageous in terms of the engine thermalefficiency.

Preferably, in the method of the present invention, throughout theengine operating region subject to the compression self-ignitioncombustion, an excess air ratio λ which is a ratio of an actual air-fuelratio to the stoichiometric air-fuel ratio, is set to 2 or more.

Preferably, in the apparatus of the present invention, the controller isoperable, throughout the engine operating region subject to thecompression self-ignition combustion, to set an excess air ratio λ whichis a ratio of an actual air-fuel ratio to the stoichiometric air-fuelratio, to 2 or more.

In the above method and apparatus, the compression self-ignitioncombustion is performed under a significantly lean air-fuel ratiocorresponding to an excess air ratio λ of 2 or more. This has anadvantage of being able to effectively reduce an amount of NOx to beproduced from combustion so as to sufficiently meet emissionregulations.

This application is based on Japanese Patent Application Serial No.201-014907, filed in Japan Patent Office on Jan. 27, 2010, the contentsof which are hereby incorporated by reference.

Although the present invention has been fully described by way ofexample with reference to the accompanying drawings, it is to beunderstood that various changes and modifications will be apparent tothose skilled in the art. Therefore, unless otherwise such changes andmodifications depart from the scope of the present invention hereinafterdefined, they should be construed as being included therein.

What is claimed is:
 1. A method for controlling an engine, comprisingsteps of: allowing a compression self-ignition combustion under anair-fuel ratio leaner than a stoichiometric air-fuel ratio to beperformed at least in a partial-load range of the engine, and acquiringa compression end temperature which is an in-cylinder temperature justbefore an air-fuel mixture self-ignites, wherein, under a condition thatan engine speed varies at a same load in an engine operating region ofthe compression self-ignition combustion, the compression endtemperature is controlled to be raised higher in a higher engine speedside than in a lower engine speed side.
 2. The method as defined inclaim 1, wherein the control for raising the compression end temperatureincludes a control for raising a compression initial temperature whichis an in-cylinder temperature at a start timing of a compression stroke,and wherein, the compression initial temperature is controlled to beraised higher in the higher engine speed side than in the lower enginespeed side.
 3. The method as defined in claim 2, wherein the control forraising the compression initial temperature includes a control forperforming internal EGR for causing burned gas to remain in a cylinder,and wherein, an internal EGR amount is controlled to be increased largerin the higher engine speed side than in the lower engine speed side. 4.The method as defined in claim 3, wherein, when the internal EGR isperformed at a same engine speed, the internal EGR amount is increasedas the load becomes lower.
 5. The method as defined in claim 2, furthercomprising a step of raising a compression initial pressure which is anin-cylinder pressure at a start timing of a compression stroke,according to a rise in the compression initial temperature, by means ofsupercharging.
 6. The method as defined in claim 1, further comprising astep of setting a geometric compression ratio of the engine to 16 ormore.
 7. The method as defined in claim 1, wherein, throughout theengine operating region subject to the compression self-ignitioncombustion, an excess air ratio λ which is a ratio of an actual air-fuelratio to the stoichiometric air-fuel ratio, is set to 2 or more.
 8. Anapparatus for controlling an engine, comprising: a controller adapted tocontrol respective sections of the engine to allow a compressionself-ignition combustion under an air-fuel ratio leaner than astoichiometric air-fuel ratio to be performed at least in a partial-loadrange of the engine, wherein the controller is operable, under acondition that an engine speed varies at a same load in an engineoperating region of the compression self-ignition combustion, to controla compression end temperature, which is an in-cylinder temperature justbefore an air-fuel mixture self-ignites, in such a manner that it israised higher in a higher engine speed side than in a lower engine speedside.
 9. The apparatus as defined in claim 8, wherein the control forraising the compression end temperature includes a control for raising acompression initial temperature which is an in-cylinder temperature at astart timing of a compression stroke, and wherein the controller isoperable to control the compression initial temperature in such a mannerthat it is raised higher in the higher engine speed side than in thelower engine speed side.
 10. The apparatus as defined in claim 9,wherein the control for raising the compression initial temperatureincludes a control for performing internal EGR for causing burned gas toremain in a cylinder, and wherein the controller is operable to controlan internal EGR amount in such a manner that it is increased larger inthe higher engine speed side than in the lower engine speed side. 11.The apparatus as defined in claim 10, wherein the controller isoperable, when the internal EGR is performed at a same engine speed, toincrease the internal EGR amount as the load becomes lower.
 12. Theapparatus as defined in claim 9, wherein the engine is equipped with asupercharger operable to supercharge fresh air, and wherein thecontroller is operable to activate the supercharger to raise acompression initial pressure which is an in-cylinder pressure at a starttiming of a compression stroke, according to a rise in the compressioninitial temperature.
 13. The apparatus as defined in claim 8, wherein ageometric compression ratio of the engine is set to 16 or more.
 14. Theapparatus as defined in claim 8, wherein the controller is operable,throughout the engine operating region subject to the compressionself-ignition combustion, to set an excess air ratio λ which is a ratioof an actual air-fuel ratio to the stoichiometric air-fuel ratio, to 2or more.